1. Field of the Invention
The present invention relates to a fuel injection control system for an internal combustion engine of cylinder injection type (also called direct injection type), which system is designed to inject directly a fuel into a plurality of cylinder chambers of the engine for allowing a fuel mixture to be burnt or combusted in a lean state in response to spark ignition. More particularly, the invention is concerned with a fuel injection control system for an engine of cylinder injection type which system is so arranged as to suppress occurrence of the so-called torque shock upon changeover of combustion modes while protecting combustion performance from degradation to thereby enhance drivability of a motor vehicle equipped with the engine.
2. Description of Related Art
Heretofore, there has been well known in the art such a cylinder injection type internal combustion engine in which the fuel injectors are disposed in the individual cylinders, respectively, for injecting directly the fuel into the combustion chambers as well as the fuel injection control system therefor. By way of example, reference is to be made to Japanese Patent Application Laid-Open No. 312396/1996 (JP-A-8-312396).
In the cylinder injection type internal combustion engine, torque generated by the engine will change in dependence on the air-fuel ratio even in the state where the throttle valve is held at a same opening degree. Accordingly, it is necessary to set optimally the combustion parameters such as the ignition timing and the fuel injection timing in dependence on the engine load and the air-fuel ratio by controlling properly the opening degree of the throttle valve and hence the air-fuel ratio.
For having better understanding of the principle underlying the present invention, technical background thereof will be described below in some detail. FIG. 12 is a schematic diagram showing generally an arrangement of a conventional fuel injection control system for a cylinder injection type internal combustion engine known heretofore.
Referring to FIG. 12, an internal combustion engine (hereinafter also referred to as the engine) 1 is equipped with an intake pipe 1a for introducing the intake air into the engine and an exhaust pipe 1b for discharging the exhaust gas resulting from the combustion of the air-fuel mixture.
An air flow sensor 2 for detecting a flow rate or quantity Qa of the intake air fed to the engine 1 as indicated by an arrow is installed at an upstream location in the intake pipe 1a.
Further installed within the intake pipe 1a is a throttle valve 3 for adjusting or regulating the intake air flow rate Qa, wherein a throttle position sensor 4 for detecting the opening degree .theta. of the throttle valve 3 is provided in association therewith.
Installed at a downstream location within the intake pipe 1a, e.g., at a location immediately preceding to the engine 1 is a surge tank 5.
On the other hand, an air-fuel ratio sensor 6 which may be constituted by a linear type O.sub.2 -sensor is provided in association with the exhaust pipe 1b for detecting an actual air-fuel ratio F of the exhaust gas, which ratio ordinarily lies within a range of e.g. 10 to 50.
A throttle valve actuator 7 is provided in association with the throttle valve 3 for adjusting the opening degree .theta. thereof. This actuator 7 may be constituted, for example, by a stepping or stepper motor which is designed for stepwise driving rotationally the throttle valve 3 to thereby adjust the rate or quantity of the intake air flowing through the intake pipe 1a.
Installed within each of the cylinders of the engine 1 is a spark plug 8 at which electric spark discharge is caused to take place for igniting the air-fuel mixture charged into the combustion chamber defined within the cylinder. To this end, a distributor 9 is provided for supplying a high voltage distributively to the individual spark plugs 8 in conformance with proper ignition timing.
Further installed is an ignition coil 10 which is realized in the form of a transformer having primary and secondary windings. A high voltage required for the spark ignition is induced in the secondary winding of the ignition coil 10 whenever a primary current flowing through the primary winding is interrupted. The high voltage is then supplied to the distributor 9. Provided in association with the ignition coil 10 is an ignitor 11 which is constituted by a power transistor for interrupting the current flowing through the primary winding of the ignition coil 10 in conformance with the ignition timing for the engine cylinders.
The spark plug 8, the distributor 9, the ignition coil 10 and the ignitor 11 cooperate to constitute a so-called ignition system for igniting or firing the air-fuel mixture within the individual cylinders of the engine 1.
Each of the engine cylinders is equipped with a fuel injector 13 for injecting directly the fuel into the cylinder chamber. A crank angle sensor 14 for generating a crank angle signal CA is provided in association with the crank shaft which is driven rotationally by the engine 1.
The crank angle sensor 14 is designed to output a pulse signal corresponding to the engine rotation number or engine speed (rpm) as the crank angle signal CA and serves also as an engine rotation sensor (or engine speed sensor), as is well known in the art. Further, the crank angle signal CA contains pulses having edges which represent the reference crank angles for the individual cylinders, respectively, wherein the reference crank angles are used for arithmetically determining various control timings for operation of the engine 1.
An accelerator pedal (not shown) manipulated by an operator or driver is provided with an accelerator pedal position sensor 15 for detecting the accelerator pedal stroke .alpha..
An exhaust gas recirculation passage (hereinafter also referred to as the EGR passage) 16 is provided between the exhaust pipe 1b and the surge tank 5 for the purpose of recirculating a part of the exhaust gas into the intake pipe 1a, wherein a stepping-motor-driven type EGR regulating valve 17 (constituting a part of the EGR regulating means) is provided in association with the EGR passage 16 for regulating the amount or quantity of the exhaust gas recirculated to the intake pipe. This quantity is referred to as the EGR quantity.
An ECU (Electronic Control Unit) 12 which is in charge of controlling the engine system as a whole is comprised of a microcomputer for arithmetically determining control quantities or parameters for various actuators which are installed for controlling the fuel combustion in the engine 1 on the basis of information detected by various types of sensors (i.e., information concerning the operation states of the engine 1), to thereby issue driving signals indicative of control quantities to the relevant actuators.
As the control signals, there may be mentioned an intake-air quantity control signal A for the throttle actuator 7, an ignition timing signal G for the ignitor 11 (and hence for the ignition system), an injection pulse signal J for each of the fuel injector 13, an EGR control signal E for the EGR regulating valve 17 and others.
As other sensors not shown in the drawings, there may be mentioned an intake air pressure sensor disposed within the intake manifold of the engine for detecting the intake air pressure of the engine 1 (also referred to as the boost pressure Pb which represents the intra-cylinder intake air quantity), a water temperature sensor for detecting the temperature of cooling water for the engine 1 and the like.
In general, the engine 1 has different types of combustion modes which include a homogeneous combustion mode in which the fuel injection is carried out during the suction stroke and a stratified combustion mode in which the fuel injection is carried out during the compression stroke.
FIG. 13 is a view for illustrating the combustion modes (or fuel injection modes) which are set in dependence on the engine operation states represented, for example, by the engine speed Ne and the engine load.
Referring to FIG. 13, the combustion modes are changed over sequentially as the engine speed Ne and the engine load (represented by the accelerator pedal stroke .alpha. or the intake air quantity information Qa or the like) increase. More specifically, starting from a so-called compression lean mode (i.e., combustion mode in which fuel is injected during the compression stroke in such an amount that a lean air-fuel mixture prevails within the cylinder), the combustion mode can be changed over to a suction lean mode (i.e., combustion mode in which the fuel is injected during the suction stroke in such an amount that a lean air-fuel mixture prevails within the cylinder) and hence to a stoichiometric ratio feedback mode (i.e., combustion mode in which the fuel is injected during the suction stroke in such an amount that the air-fuel mixture of the stoichiometric air-fuel ratio prevails within the cylinder) and hence to an open loop mode (i.e., combustion mode in which the fuel injection quantity is increased without validating the feedback control).
In the compression lean mode, combustion is realized with extremely lean fuel mixture because of the fuel injection carried out during the compression stroke.
Further, in the suction lean mode, combustion is realized with a lean fuel mixture (with a greater air-fuel ratio when compared with the stoichiometric ratio) due to the fuel injection during the suction stroke even though the fuel mixture is not so lean as in the compression lean mode.
Furthermore, in the stoichiometric ratio feedback mode, combustion is effected with the stoichiometric air-fuel ratio on the basis of the oxygen concentration signal derived from the output of the air-fuel ratio sensor.
Finally, in the open loop mode, combustion is effected in the fuel-rich state without validating the feedback control.
With the fuel injection control system for the cylinder injection type internal combustion engine described above, not only the engine output performance but also the fuel consumption can be improved by changing over the target or desired air-fuel ratio between the stoichiometric air-fuel ratio on the order of 14.7 and the large air-fuel ratio on the order of 20 to 30 (lean mixture) in dependence on the combustion modes.
When the desired air-fuel ratio is to be changed in dependence on the combustion modes (i.e., upon changeover of the combustion modes), the fuel injection quantity which can be represented by a combination of the air-fuel ratio A/F and the cylinder-charged intake air quantity Qai is so adjusted that the available engine output torque can remain unchanged for the accelerator pedal stroke .alpha. which reflects the driver's intention in order to suppress the torque shock.
FIG. 14 is a view for illustrating the output torque characteristics (obtained experimentally) of the engine 1. More specifically, shown at (a) in FIG. 14 is the engine output torque characteristics observed when the boost pressure Pb equivalent to the cylinder-charged intake air quantity Qai is changed with the fuel injection quantity Qf being maintained constant while illustrated at (b) in FIG. 14 is the engine output torque characteristic observed when the fuel injection quantity Qf is changed with the boost pressure Pb, i.e., the cylinder-charged intake air quantity, being maintained constant.
In the case illustrated at (a) in FIG. 14, the pumping loss decreases in proportion to the increase of the cylinder-charged intake air quantity Qai. Thus, the output torque of the engine 1 bears a proportional relation to the cylinder-charged intake air quantity Qai.
On the other hand, in the case illustrated at (b) in FIG. 14, the output torque of the engine 1 bears a proportional relation to the fuel injection quantity Qf.
Parenthetically, in FIGS. 14(a) and (b), the point a represents the torque value in the stoichiometric mode while the point b represents the torque value in the compression lean mode (with a large air-fuel ratio (lean mixture)). By way of example, in a steady running state succeeding to the acceleration, the stoichiometric air-fuel ratio (based on the point a) is changed over to a large or lean air-fuel ratio (based on the point b).
When the combustion mode is changed over from the stoichiometric mode to the compression lean mode with the throttle valve opening degree a being increased to thereby allow the boost pressure Pb (or cylinder-charged intake air quantity Qai) to increase from a level corresponding to the stoichiometric air-fuel ratio value Pba to a level corresponding to the lean air-fuel ratio value Pbb, the output torque will increase by a quantity .DELTA.Tp due to decrease of the pumping loss (see FIG. 14 at (a)).
Consequently, even when the fuel injection quantity Qf is held constant after the changeover of the combustion mode, there takes place a torque difference corresponding to the above-mentioned torque change .DELTA.Tp even in the state where the accelerator pedal stroke is held constant by the driver.
Such being the circumstances, it becomes necessary to generate a change quantity .DELTA.Tf (see FIG. 14 at (b)) by adjusting the fuel injection quantity Qf so as to cancel out the torque difference, i.e., the torque change quantity .DELTA.Tp mentioned above.
By way of example, in the conventional system described in Japanese Patent Application Laid-Open No. 312396/1996 (JP-A-8-312396) mentioned hereinbefore, the air-fuel ratio to be controlled is changed by varying linearly the reciprocal F/Ao of the desired air-fuel ratio A/Fo upon changeover of the combustion modes.
More specifically, for changing over the combustion mode to the lean side, the reciprocal F/Ao of the desired air-fuel ratio is changed linearly in the leaning direction (i.e., decreasing direction).
FIG. 15 is a timing chart for illustrating the combustion mode changeover operations of the conventional system disclosed in the publication mentioned above. In the figure, the time base for the control cycles of the engine 1 is taken along the abscissa.
More specifically, shown in FIG. 15 is a steady running state of the engine in which the combustion mode flag is changed over to the compression lean mode (i.e., the stratified combustion mode) from the stoichiometric mode (i.e., combustion mode with the stoichiometric air-fuel ratio).
As can be seen from FIG. 15, when the desired throttle valve opening degree .theta.o is changed so as to increase upon changeover of the combustion modes, then the actual throttle valve opening degree .theta. is also changed stepwise instantaneously in the increasing direction in response to the changeover of the desired throttle valve opening degree .theta.o.
On the other hand, the boost pressure Pb (cylinder-charged intake air quantity Qai) increases along a curve representing a delay from the changeover time point, as can be seen in FIG. 15. This is because the intake air fed through the intake pipe can reach the engine 1 with a delay after having been stored in the surge tank 5 and the intake manifold, respectively. This delay is referred to as the first-order delay for convenience of the description.
Thus, it is apparent that the cylinder-charged intake air quantity Qai which increases with the first-order delay, as described above, does not coincide with the reciprocal F/Ao of the desired air-fuel ratio which changes linearly. Consequently, the fuel injection quantity Qf which is so controlled as to decrease as a function of the reciprocal F/Ao of the desired air-fuel ratio will undergo inevitably such a variation as illustrated in FIG. 15 under the influence of the first-order delay of the change of the cylinder-charged intake air quantity Qai. As a result of this, in the transient state intervening the combustion modes changed over, the torque change quantity .DELTA.Tp ascribable to the increase of the boost pressure Pb (cylinder-charged intake air quantity Qai) can not be canceled out by the torque change quantity .DELTA.Tf brought about by the control of decreasing the fuel injection quantity Qf, which results in variation of the overall torque change quantity .DELTA.Tr (=Tp+.DELTA.Tf).
Parenthetically, FIG. 15 shows the behaviors of the engine in the course of mode changeover from the stoichiometric mode to the compression lean mode. It should however be understood that similar torque shock may take place in the mode changeover from the compression lean mode to the stoichiometric mode.
As is apparent from the foregoing, the conventional fuel injection control system for the cylinder injection type internal combustion engine suffers a problem that the torque change .DELTA.Tf brought about by the change of the fuel injection quantity Qf can not cancel out the torque change .DELTA.Tp ascribable to the change of the boost pressure Pb (cylinder-charged intake air quantity Qai) because in the transient state which accompanies the changeover of the combustion mode, the desired air-fuel ratio is changed linearly for the boost pressure Pb (cylinder-charged intake air quantity Qai) which changes with the first-order delay. Thus, occurrence of the torque shock can not be suppressed, giving rise to a problem.